1. Field of the Invention
This invention relates to a starter motor with an intermediate gear in which the rotation of a pinion disposed externally of a clutch inner of an overrunning clutch mounted on an output shaft of the starter motor is reduced and transmitted through an intermediate gear to a ring gear.
2. Description of the Prior Art
In FIG. 2, there is shown a section view of a conventional starter motor with an intermediate gear. In this figure, reference numeral 1 designates a DC motor which comprises a stator 2 and an armature 3. 4 stands for an output shaft which is an extension of an output shaft of the armature 3, and 5 points out a front bracket which is coupled to the stator 2 by means of a bolt and supports the front end portion of the output shaft 4 through a bearing 6. 7 designates an end cover. 8 stands for an overrunning clutch which is mounted to the output shaft 4 and constructed in the following manner: that is, the overrunning clutch 8 includes a clutch outer member 9 coupled to the output shaft 4 through a helical spline 11, and a clutch inner member 10 which transmits rotation to the clutch outer member 9 in one way through a roller 12. The clutch inner member 10 includes a pinion 13 in the front end portion thereof. The clutch inner member 10 is supported by the output shaft 4 through a bearing 15, and is slidable in the axial direction thereof.
16 designates a stopper which is fixed to the output shaft 4 and is used to receive the advancing movement of the overrunning clutch 8 at a given position. 17 stands for a shift lever which is rotatably supported in the intermediate portion thereof by the front bracket 5, with the upper end portion of the shift lever 17 in engagement with a plunger (not shown) of an electromagnetic switch. The shift lever 17 also includes a forked portion the lower end portion of which is in axial engagement with an annular engaging portion 14 of the clutch outer member 9. When the electromagnetic switch is electrically energized, the shift lever 17 is rotated counterclockwise in FIG. 2 to thereby move the overrunning clutch 8 forwardly. When the electromagnetic switch is cut off, then the shift lever 17 is moved or returned clockwise in FIG. 2 to thereby move backwardly or return the overrunning clutch 8 to its original position.
18 designates an intermediate gear shaft fixed to the front bracket and 19 stands for an intermediate gear which is supported rotatably and axially movably by the intermediate gear shaft 18 through a bearing 20. The intermediate gear 19 is always in mesh with the pinion 13 and is rotatable reducingly. 21 points out an annular linking member which is fixedly secured to the outer peripheries of the clutch outer member 9 and includes a flange portion 21a in engagement with an annular groove 19a formed in the intermediate gear 19. The annular linking member 21 moves the intermediate gear 19 in link with the forward or rearward movement of the overrunning clutch 8. 25 designates a ring gear disposed on a flywheel of an engine. The ring gear 25 is engageable with the intermediate gear 19 when the latter is moved forwardly, so that the ring gear 25 can be started and rotated.
Next, description will be given below of the engagement and shift of a set of gears with reference to FIG. 4 in which a tooth 30a of a small gear 30 is in mesh or engagement with a tooth 31a of a large gear 31.
Assuming that m=a module, Z=the number of teeth; dp=Z.times.m=pitch diameter; X=a shift coefficient; Xm=an amount of shift; .alpha.c=a tool pressure angle; dg=a base circle diameter; a=a center distance; Cn=a backlash; and .alpha.b=an engaging pressure angle, then the following equation can be obtained. EQU cos.alpha.b=(dp.sub.1 +dp.sub.2).multidot.cos.alpha.c/(2a) (1)
Profile shifted gears are gears which are formed by shifting the standard pitch line (a pitch line for which the tooth thickness is one half of a pitch) of a rack tool in the radial direction from the standard pitch circle radii (r.sub.0 =d.sub.p /2=Zm/2) of the gears. This can be accomplished by cutting the teeth of a gear with an enlarged or reduced outside diameter, thereby increasing or decreasing the tooth thickness, respectively. In standard gears, when the number of teeth is small, then undercutting occurs to reduce the value of the meshing or action rate of the gears. That is, in the set of gears, the teeth of the small gear is weaker in strength than those of the large gear. For this reason, if the gears are formed as the profile shifted gears, then various kinds of requirements can be satisfied. For example, the undercutting can be prevented according to the designs of the gears by use of a standard tool, the tooth thickness and center distance of the gears can be varied, and so on. The degree by how much tooth thickness is varied is the shift coefficient. The higher the shift coefficient, the greater the tooth thickness, and accordingly, the greater the strength of the teeth.
In a gear arrangement in which a pair of gears are included, the number of teeth increases sequentially in the order of the pinion 13, intermediate gear 18 and ring gear 25. For this reason, in view of the gear strength, conventionally, the pinion 13 is given the greatest shift coefficient, the intermediate gear 18 the intermediate shift coefficient, and the ring gear 25 the smallest shift coefficient.
Now, description will be given below of the engagement relations among the pinion 13, intermediate gear 18 and ring gear 25 employed in a conventional intermediate gear type starter motor with reference to FIG. 3(A).
Assuming that .alpha.b.sub.1 =an engaging pressure angle between the pinion 13 and intermediate gear 19, and .alpha.b.sub.2 =an engaging pressure angle between the intermediate gear 19 and ring gear 25, then the shift coefficient x of the pinion&gt;the shift coefficient x of the intermediate gear&gt;the shift coefficient x of the ring gear. Thus, .alpha.b.sub.1 is considerably greater than .alpha.b.sub.2. If a difference between .alpha.b.sub.1 and .alpha.b.sub.2 is great, then the intermediate gear 19 is given pressure and is drawn to one side due to the difference between the two angles, thereby worsening the engagements therebetween, as shown in FIG. 3(B). In FIG. 3(B), F.sub.1 expresses an action force to be given to the intermediate gear 19 by the pinion 13, F.sub.2 stands for a reaction force to be produced in the pinion 13 from the intermediate gear 19, F.sub.3 points out an action force to be given to the ring gear 25 by the intermediate gear 19, and F.sub.4 designates a reaction force to be produced in the intermediate gear 19 from the ring gear 25. While the respective forces are all equal, the engaging pressure angles .alpha.b are different. Also, an engagement efficiency obtained between the pinion 13 and intermediate gear 19 is worsened (because the engaging pressure angle .alpha.b is great). When the state of FIG. 3(B) is viewed from the whole intermediate gear 19, then the component of a composite force P produced by F.sub.1 and F.sub.4 becomes greater, resulting in the biased engagement, as shown in FIG. 3(C). o Here, description will be given of an engagement coefficient between a pair of gears with reference to FIG. 5. The term "engagement coefficient" means that pieces of teeth are engaged when a pair of gears are in engagement with each other. The engagement coefficient .epsilon.=the engagement length/the normal pitch. As the engagement coefficient varies, the loads to be applied to the teeth of the gears are caused to vary. In theory, when the engagement coefficient is equal to 1 or less, then the gears cannot be rotated in a normal condition. The engagement coefficient .epsilon. is expressed by an expression (2) of a numerical representation 1. Assume that dg=bottom land circle diameter, and dk=tooth crest circle diameter. ##EQU1##
Normal pitch (te): a distance between the tooth surfaces measured at right angles to the tooth surfaces ##EQU2##
(.alpha.b expresses the engagement pressure angle)
(With respect to respective designations, see FIGS. 4, 5 and the description from line 19 of page 2 to line 8 of page 3 in the specification. For example ST represents the length of the line connecting points S and T in FIG. 5, MN represents the length of the line connecting points M and N, etc.)
From the equation (2) of the numerical representation 1, it is found that the engagement coefficient .epsilon. is great when the engagement pressure angle .alpha.b is small, and it is small when .alpha.b is great.
In general, the number of the teeth of a ring gear employed in an engine is of the order of 100 and the number of the teeth of a pinion in the engine is of the order of 8 to 10. When only the pinion and ring gear are employed, the shift of the ring gear is set small and the shift of the pinion is set large, thereby balancing the strength of the teeth. Also, since the engagement pressure angle .alpha.b can be obtained according to the above-mentioned equation (1), a shift on one side is small and a shift on the other side is great, so that the denominator 2a (a: center distance) of the equation (1) does not become too large. Therefore, the engagement pressure angle .alpha.b does not become too large but provides a suitable value. As a result of this, a relatively large engagement coefficient can be obtained. Assuming that the intermediate gear 19 is not changed, if the above-mentioned relation is maintained between the intermediate gear 19 and ring gear 25, been both of the intermediate gear 19 and ring gear 25 provide large shifts to increase the denominator 2a, with the result that the engagement pressure angle .alpha.b also becomes large. In other words, while the engagement pressure angle .alpha.b between the intermediate gear 19 and ring gear 25 is small and the engagement coefficient thereof is large, the engagement pressure angle between the pinion 13 and the intermediate gear 19, is large and the engagement coefficient thereof is small.
In the above-mentioned conventional starter motor with an intermediate gear, a force acting on the intermediate gear 19 operates in such a manner that the engagement pressure angle between the pinion 13 and intermediate gear 19 is considerably larger than the engagement pressure angle between the intermediate gear 19 and ring gear 25. For this reason, the radial loads of these two engagement pressure angles are not balanced with each other, whereby the intermediate gear 19 is caused to come nearer by the clearance with respect to the intermediate gear shaft 18. As a result of this, an expected engagement cannot be obtained. For example, a backlash is large on one side while it is small on the other side, with the result that there can be produced a strange sound and a bearing (sleeve bearing) 20 of the intermediate gear 19 can be worn too excessively. Also, as described above, due to the fact that engagement coefficient between the pinion 13 and intermediate gear 19 is lowered, the teeth thereof are worn to a greater extent.